A multistage evaporation organic rankine cycle

ABSTRACT

The invention ECT relates to methods for improving the amount of electricity gained from preferably waste heat by a normal or an organic (ORC) Rankine process with vaporization in several stages, normally three. The waste heat in sensible form is exchanged in at least two in series coupled evaporators to a receiving working fluid (e.g. a refrigerant) that passes at least two of said evaporators, but coupled in parallel. Of the waste heat between the temperature of the heat source and that of the heat sink about 80% can be used for direct electricity generation. An embodiment of the invention uses a radial turbine with a centripetal (inwards) flow direction. The different vapor enthalpies from the said vaporization stages enters a turbine wheel/runner 51 at different outside diameters D2, D2′ and/or with suitable tangential velocities obtained by different guiding vane sets 65, 66 and 67.

TECHNICAL FIELD

The invention relates to an apparatus of Organic Rankine Cycle type including a closed loop working fluid circuit operable between a heat source and a heat sink. More specifically the invention relates to ORC apparatuses having a working fluid circuit including: a heat exchanger arrangement for vaporizing and/or superheating a working fluid by exchanging energy from the heat source; at least one turbine for expanding the vaporized/superheated working fluid; condensing means connectable to the heat sink for condensing the expanded working fluid from the turbine; and pumping means for pumping and pressurizing the condensed working fluid to the heat exchanger arrangement.

BACKGROUND

A normal or an Organic Rankine Cycle/process (ORC) gives up to now poor results as the working media normally vaporize at a constant (isothermal) temperature. Better results are obtained by non-isothermal conditions received normally by ammonia-water mixtures, as e.g. in the Kalina process. However, these require complicated process structures and some difficulties in selecting a suitable turbine/expander especially regarding its shaft speed.

SUMMARY OF THE INVENTION

The invention relates to an improved ORC of the type initially mentioned, in which the heat exchanger arrangement comprises at least three parallel coupled evaporators forming at least three pressure stages on the working fluid side, and providing one outlet per pressure stage for connecting to the at least one turbine.

Preferably, the at least one turbine is a single turbine including one inlet per pressure stage. Each inlet connected to respective pressure stage outlet. The turbine further having a common outlet for the expanded working fluid.

In some embodiments, the turbine may drive an electric generator. In other embodiments the turbine may e.g. drive a propeller axis, a compressor or a pump.

Preferably, the heat exchanger arrangement includes three evaporators, one for a low pressure stage, one for a medium pressure stage, and one for a high pressure stage, at least one economizer per pressure stage, and optionally at least one superheater per pressure stage.

Preferably, the heat exchanger arrangement includes: one economizer at the low pressure stage, two economizers at the medium pressure stage, and three economizers at the high pressure stage. The economizers at respective pressure stage being serial connected (if more than one) on the working fluid side.

Preferably, the heat exchanger arrangement is configured to heat the working fluid at each pressure stage to an individual temperature starting point at the corresponding turbine inlet, each starting point selected to be in the dry/superheated region. Preferably, the heat exchanger arrangement is configured such that each starting point is selected to provide the turbine expansion to end in within +_5° C. of a common temperature and pressure end point in the dry/superheated region. The common end point is preferably situated about 5-10% of the latent heat/enthalpy in the dry/superheated region of the saturation curve from the saturation curve. More preferably, the common end point is situated about 1-2% of the latent heat/enthalpy in the dry/superheated region of the saturation curve from the saturation curve.

Preferably, at least two of the evaporators are coupled in series on the heat source side, such that at least a portion of the heat source flow is directed through the evaporators in series. More, preferably at least three evaporators are coupled in series on the heat source side.

In one embodiment, the pressure stages are selected such that the vaporization temperatures of the evaporators are within T2+A_(n)*(T1−T2)+−15%, where A_(n)=(1/(n+1), 2/(n+1), . . . , (n)/(n+1)), where n being the number of evaporators, and where T1 being the temperature of the heat source and T2 the temperature of the heat sink. Preferably the variation is selected as +−10%.

A new turbine design matching the new heat arrangements is suggested. The turbine is of a turbo machine type, e.g. an impulse/action type or a reaction turbine and having at least two inlets, and directing means in the form of guiding vanes and/or nozzles for directing a flow from said inlets on to different radii of one common turbine wheel of the turbine.

The turbine may be a reaction turbine preferably of a radial or a mixed flow type with a centripetal (inwards) flow direction with a casing and at least one runner/wheel with working blades, at least two inlets and one common outlet where the inlets in turn are connectable to guiding vanes/nozzles in the turbine casing for expansion of the gas at different radii.

The turbine may alternatively be an impulse/action type with a disc that has blades at different radii to match the driving expanded flow from the nozzles whereby at least two sets of the nozzles are situated at different radii and the remaining at equal radii but at separate angular sectors. Preferably having at least two of the said guiding vanes/nozzles situated on different diameters D4 and D4′ respectively. Preferably, having the said runner working blades with an extension to at least two different diameters D2 and D2′ respectively.

The turbine/expander may be of a positive displacement machine type with several stages by having displacement volumes determined by a first actual flow stream of the working fluid to which part working fluid streams are further added between the stages and that the downstream displacement sizes are correspondingly increased to swallow the total vapor flows. Preferably, displacement machine is a screw expander/turbine with one or several screws designed to be a 2 or 3 stage machine with intermediate inlets for said part flow streams between the stages and that the downstream screw displacement/volume capacity is designed by selecting shape, size and/or the number of the “pistons” in a first/male and slots in a second/female rotor cooperating with the first to swallow both the actual and the total volume flows.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will further down be presented in detail with reference to the drawings/figures in which:

FIG. 1 Shows waste heat power process circuits of prior art.

FIG. 2 shows the temperature profile with T as temperature and Was heat flow.

FIG. 3 Legend: Case 70 degrees: thin black and dashed black (=heat source). Optimal for power output thick black and dashed thick black (heat source). Case with 13 degrees below heat source as claimed by many to be optimal: thin black and dashed black (=heat source).

FIGS. 3a and 3b shows the temperature profile for the heat source when heat is removed at a high and a low temperature range respectively.

FIG. 4 Individual temperature profiles

FIG. 5 The temperature profile−temperature (y-axis)/heat flow (x-axis) curves

FIG. 6 Circuits for the working fluid.

FIG. 7 Alternative for the circuits for the working fluid.

FIG. 8 Heat source circuit. Figures “13H”, “13M” and “13L” mark the different vaporization stages.

FIG. 8a Alternative for heat source circuit.

FIG. 9 Diagram log(p) versus enthalpy h (p=absolute pressure and h enthalpy

FIG. 10 Axial turbine efficiency at different velocity ratios u/c.

FIG. 11 Section of an axial turbine wheel (left) and inlet vanes/nozzles (right).

FIG. 12 Section of a radial turbine for three different admission data. Circuits in the volutes 61 and 62 are obtained by dividing the peripheral circumference in two parts.

FIG. 13 Alternative for the circuits for the working fluid.

FIG. 14 Alternative for heat source circuit.

STATE OF THE ART Process Circuit

A waste heat process using sensible heat has normally three different circuits; a heat source circuit 1, a working fluid circuit 2 and a heat sink 3, FIG. 1.

The fluid of the heat source (gas or liquid) passes and exchanges heat, in a possible superheater 14, an evaporator 13 and an economizer 12 with a decrease in temperature from 1′ to 1″.

The working fluid in 2 passes said heat exchangers with a change in state from liquid to vapor with a corresponding increase of its energy content from 2 to 2″. In a turbine/expander 15 connected to an electric generator 159 the fluid expands and gives away energy.

The vapor of fluid 2 is led to a condenser 16 and from this to a possible tank 17 preferably placed below the condenser. A feed pump 11 closes the circuit by pumping the fluid into said heat exchangers.

The heat sink 3 cools the condenser and may have a pump 17. The heat sink may e.g. be ambient (cold) water or air or indirectly a liquid from a cooling tower or an air-cooled radiator.

Temperature Profile

A graph, FIG. 2, with heat flow Was abscissa (x-axis) and temperature T as ordinate (y-axis) gives an excellent view of all important data.

The heat source 1 is given by an approximately straight line, the working process 2 by a closed circuit 2 and the heat sink by the line 3. The said components are marked with equal figures as in FIG. 1.

In the circuit 2 the evaporation is shown as the line 13, the economizer by 12 and the superheater by 14. The line 13 is shown oblique to the W-axis that means the liquid vaporizes at varying temperatures or at non-isothermal conditions (as in the Kalina process). This is typical for mixtures of different media. For a pure fluid or an azeotrophic mixture the line 13 is horizontal as it vaporizes at a constant temperature. Similar conditions are valid for the line 16 at the condenser.

An expansion of the working fluid is marked 15 and 151 of which 15 represents the useful mechanical work obtainable as input/output to/from the turbine. The size of this work is also given as the distance 0-150 while the total source heat flow is 0-1″.

The temperature profile as shown in FIG. 2 is split in two parts an outer or gross part defined from the heat source 1 and heat sink 3 with a maximal temperature difference from the points 1′ and 1″ with temperatures T₁ and T₂ respectively and the difference T. These corresponding temperatures give a gross efficiency.

An inner or net part is defined by the line 2 and 16 with a temperature difference t1−t2. From the geometry it is obvious that these two are given solely by 14″, 12″ and 16′. The “knee” at 12″ is in English literature called “pinch point”. A net efficiency is obtained from t1 and t2 with the difference t.

For the said heat exchangers including the condenser, temperature differences are required to drive a heat transfer. In FIG. 7 these differences are marked 12′, 12″, 13″, 14″ and 16′. The size of the heat exchanger areas depends on the max and the min difference. An optimal area sizing can now be based on the said gross efficiency and the said temperature differences from the temperature profile.

A temperature heat flow part marked as 151 is the waste heat from the inner process 2. As the temperature is above that of the heat sink it can, known per se, be recovered by a recuperator and utilized for an improvement of the inner process shown in FIG. 2 as 152, parallel moved to the right as a tail for increasing the heat source.

Prior Art General Methods Layout.

Known, elementary and rather common is to use evaporation at 2 different temperatures (pressures) especially in combination with several heat sources with considerably different temperatures, e.g. 250 and up and 90 degree C. respectively. Some examples are shown below with a broad variation in ways to couple two sets of components together.

U.S. Pat. No. 6,857,268 B2, a cascade solution where a low pressure circuit is fed by vapor heated by expanded vapor from a high pressure circuit. Each circuit has a separate turbine. The solution is thermodynamically poor and is possibly selected due to turbine limitations.

U.S. Pat. No. 8,438,849 B2, has two heat sources with different temperatures/pressures, where in two alternatives vapor from a high pressure turbine in series with vapor from the other sources is fed to a low pressure turbine. In a third alternative the two heat sources are coupled in parallel to independent vaporizers and independent turbines and then to a common condenser. In a forth alternative the two heat sources have independent circuits including turbines and condensers with water as working fluid for one of the circuits.

U.S. Pat. No. 8,474,262 B2, has one heat source and two separated independent circuits complete with all components as well as turbines, where the split between the two circuits is optimized.

US 2010/0071368 A1, has an ORC with a cascade coupling where expanded vapor from the high pressure turbine is fed to the low pressure turbine and in an alternative parallel circuits each coupled to a turbine. In a further alternative there are two totally independent circuits of which one has water as working fluid.

US 2010/0242476 A1, this is similar to U.S. Pat. No. 6,857,268 B2 with a cascade solution for high and low pressures circuits.

US 2010/0242479 A1, similar to US 2010/0242476 A1, but with a back end for generation of heating/refrigeration added.

US 2014/0026574 A1, has 2 heat sources and sets of evaporators, screw expanders/turbines coupled in series. The screw expanders can have wet condition for the working fluid, which normally is avoided due to damage risk in normal turbines.

US 2014/0033711A1, has evaporation at different pressures/temperatures each coupled to two expanders/turbines of the screw type for wet service.

WO 2014/0211708 has several heat sources connected to separate evaporators, turbines in parallel for the heat source fluid as well as the working fluid. The circuits are connected to one common condenser.

The present invention ECT uses in contrast to the prior art a single heat source and multiple vaporization temperature/pressures. The flow pattern through the heat exchangers (evaporators) is parallel for the working fluid and series for the heat source fluid. The reason is to gain more electric output from the same heat source by matching the temperature profile of the heat source fluid with that for the working fluid. The tail of the heat source profile will then get a low outlet temperature close to that of the heat sink.

Some Turbine Designs of the Prior Art

WO 2013/171685 A1 shows a “multistage” radial turbine design for adding/removal of a part fluid flow (intermediate superheating) with a pressure between the pressures and with adding/removal also between the normal in- and outlets. The main flow direction is centrifugal (outwards). The turbine design is similar to a Ljungström radial turbine but with one of the counter rotating part replaced by a stationary vane set.

DE 10 2012 021 357 A1 2014.05.08 shows an ORC with evaporation in two stages with heat taken from the heat source and then in a further stage with heat from a recuperator. The main vapor streams and part streams are fed into a multistage axial turbine between the turbine stages. In one of the claims the number of evaporation stages may be increased with the same principles that among other things mean that the said coupling of the heat exchangers remain and are equal.

CN 103195519 A shows an ORC with 4 evaporation stages coupled in series and driven by heat from the heat source and a further stage driven by heat from a recuperator. Vapor from the working fluid is schematically supplied to different parts of a turbine. If the working circuits are coupled in series as in this prior art separating means must be arranged between the stages as a pressure difference between liquid out from a preheater and into an evaporator due to obtain sufficient NPSH for avoiding cavitation in the pumps. Alternatively the separation could be done by a vapor/liquid tank/boiler or the pumps placed well below the exchangers. Another disadvantage is that separate pumps are necessary

DESCRIPTION OF THE INVENTION

ECT is an acronym for “Enthalpy Compounding Technology”. ECT is an invention with the intention to improve the economic conditions for generating electricity out from waste heat sources. Typical for these are that the heat is found in a sensible form that means that the temperature is gradually decreasing when heat is taken out.

The ECT Process

The ECT may take advantage of the temperature profile for a sensible heat source in that a performance optimum is selected.

According to an embodiment, the performance optimum for a process with one single evaporation process below a temperature difference below 100° C. is an average vaporization temperature (e.g. a vertical mean of line 13 in FIG. 2)) of about T2+½ *(T2−T1) with a variation of not more than +−15%. T1 being the temperature of the heat source and T2 the temperature of the heat sink.

The total or gross efficiency/gain is first and foremost defined as the net output as mechanical work in relation to the available heat between the heat source max temperature down to the minimum temperature of the heat sink. A part efficiency is sometimes erroneously calculated for the working circuit itself that, however, cannot be used for any technical and/or economical consideration.

For a simple case with the heat source line 1 as a straight line together with several evaporation temperatures the optimum will get about the same average as stated above. The FIG. 3a and FIG. 3 b shows the temperature profiles, when heat is removed from the heat source fluid. FIG. 3a shows, when heat is removed in a high temperature range and 3 b ditto in a low temperature range. The removed heat may be used for e.g. some chemical process or for the low range for district heating. The optimal value will now coincides with a gravity center for the line from 1′, Ch to 1″.

TABLE 1 An example with big differences in total efficiency in dependence of the evaporation temperature for cases, where the heat source is 90° C. and the heat sink is 20° C. The case with the average temperature, 55° C., gives the highest total efficiency. Evaporation Working circuit Total efficiency// temperature ° C. efficiency gain 77 0.1112 0.0344 70 0.1031 0.0439 55 0.0826 0.0522

In other embodiments, the ECT process may feature several different vaporization temperatures or pressure levels at which heat is transferred in the process, preferably as separate part flow circuits 2H, 2M and 2L, FIG. 4, all may have about equal temperatures t1 and t2, In FIG. 4 T is the temperature and W is the heat flow.

All the vaporization temperatures are then optimized for max output/gain. By advantage a pure or an aezotrophic fluid is used with 3 different vaporization temperatures. In the general case at least one of the pressure level may be above the critical point (critical pressure) for the working fluid, a point above which no vaporization occurs at heating. One embodiment of the invention uses basically evaporation with at least 3 different temperatures/pressures, usually 3, at at least one heat source at a certain given max temperature.

In FIG. 5 a relative temperature T_(relative) is placed on the y-axis and a relative heat W_(relative) on the x-axis. Further, the dotted line is the heat source 1 and the solid stair-shaped line represents the working fluid. The horizontal plateaus are the vaporization at the constant temperatures 2H, 2M and 2L respectively. These temperatures are optimized for highest efficiency/gain. The shown stair-shaped line gives a very good heat source utilization. However, several/infinite, infinitely small stair steps create the best possible process.

The differences at the knees (pinch points), 12″, are further economically optimized with regard to obtained electrical output/gain and the cost of the heat exchangers. The differences at 12″ are the driving “force” for the heat transfer. Typical is here differences less than 5% of the temperature span 1′ to 1″ and with preferable values of 2-3%.

The overall aim with the invention is to use the heat source as well as possible, preferable down to the heat sink. In FIG. 5 the graph shows that (1−0.82)=0.18 can't be used. The relative temperature on the waste heat in the own process then is 0.18 of the relative temperature span 1 to 0. Compared to the present conventional methods more than a doubled gain as electricity is obtained at a given and equal source of waste heat.

In an embodiment using several pressure stages the optimal vaporization temperature may be selected to be within T2+A_(n)*(T1−T2)+−15%, where A_(n)=(1/(n+1), 2/(n+1), . . . , (n)/(n+1)), where n being the number of evaporators, and where T1 being the temperature of the heat source and T2 the temperature of the heat sink. Preferably the tolerance being +−10%. E.g. for a three stage system having a heat source temp of 100° C. and a heat sink temp of 20° C., the vaporization temperatures could be selected as (40° C., 60° C., 80° C.), the values with variation of +−15%, preferably +−10%.

Heat Exchanger Arrangements

By several sets of heat exchangers the fluids flowing through them must not be equally or symmetrically coupled. In general each heat exchanger has four connection ports for in/out for the two fluids exchanging heat, compare FIG. 1. The FIGS. 6-8 show the exchanger coupling for each of the fluids, one at the time.

FIGS. 6, 7, and 13 show alternative couplings, of the heat exchanger arrangement 12-14, on the working fluid side.

ECT has according to the invention the part circuits for the working fluid high H, medium M and low L pressures in FIG. 6 coupled in parallel. I.e. the heat exchanger arrangement 12-14 comprising at least three parallel coupled evaporators forming at least three pressure stages (L, M, H) on the working fluid side. Each pressure stage connects to the turbine 15 which is designed to receive and expand vapor from each pressure stage. Pumps 11 pressurizes each pressure stage (L, M, H). The heat exchanger types 12, 13 and 14 are equal to the prior art in FIG. 1. The total solution has optimized sizing of the part flows in the circuits. For three parallel circuits the part flows are roughly one third in each. Comparing to FIG. 9, the pumps 11 raises the pressure from C00 to L0, M0, and H0 for respectively pressure stage L, M, and H. The economizers 12L, 12M, 12H preheats the working fluid to L1, M1, and H1 to respective evaporator 13L, 13M, and 13H. The evaporators 13L, 13M, 13H vaporizes the working fluid to L2, M2, and H2 to respective superheater 14L, 14M, and 14H. The superheaters 14L, 14M, and 14H superheats the working fluid to S1, S2, and S3. The superheated working fluid is then expanded to the common end point C123.

For the economizers/preheaters 12 as an alternative, some of the pipelines can be common as shown by the alternative circuit in FIG. 7. This means that all the flow passes economizer 12L and then splits for the evaporator 13L and economizer 12M and 12H. The same principle is further used for circuits H and M. Pumps 11 are for pressure rise for circuit L to M and then for M to H.

As can be seen in FIG. 13 the high and medium pressure stage economizers 12H and 12M of FIG. 6 are replaced by three 12H, 12H′ and 12H″ respectively two 12M, 12M′ economizers. These are serial coupled on the working fluid side. FIGS. 8a and 14 shows alternative couplings on the heat source side for this embodiment.

FIGS. 8, 8 a, and 14 show alternative couplings of the heat exchanger arrangement 12-14 on the heat source side.

The heat source flow circuit, FIG. 8, starting at 1′ and ending at 1″ has a main heat flow path given by the three evaporators 13H, 13M and 13L together with minor flows for the other components. The evaporators 13H, 13M and 13L are coupled in series. The reasons are to make it possible to use a maximum of heat from the heat source and simultaneously achieve small/optimal values for the pinch points at the knees, compare FIG. 5.

In the alternative heat source circuits in FIG. 8a , corresponding to the working fluid connection of FIG. 13, the basic preheaters/economizers 12H and 12M in FIG. 8 are divided in several 12M, 12M′ and 12H, 12H′ and 12H″ with connection to the heat source circuit between 13M-13L and between 13M-13L and 13H-13L respectively. The said preheaters/economizers are then heated from the heat source and its temperatures with lowest possible consumption of heat/temperature. The gain will then be more heat/temperature capacity for the downstream evaporators 13M and 13L respectively and in turn better overall heat performance.

FIG. 14, further shows an alternative couplings on the heat source side of the heat exchanger arrangement 12-14 corresponding to the working fluid connection of FIG. 13.

As can be seen in the embodiments of FIG. 8, FIG. 8a , FIG. 14 the heat exchanger arrangement of the working fluid circuits have three of the evaporators coupled in parallel, while the heat source circuits have three of them coupled in series.

Total System Design

Due to cost reduction ECT uses preferably a very small superheating of the vapor from the part evaporators normally with different values for all of them. FIG. 9 describes a pressure p enthalpy h graph. Shown is also isotherms in ° C., isenthrops (Isen) and steam/vapor quality x. The superheatings to points S1, S2 and S3 for the circuits high 2H, medium 2M and low 2L respectively, are according to the invention selected so that during and after expansion in a turbine to expanded and superheated vapor to one about common point C123.

When the working fluid has a lower slope on its saturation curve than the isenthrops corrected for turbine losses a recuperator may be used. By the common point C123, energy losses are avoided as when fluids with different temperatures (enthalpies) are mixed.

When the saturation curve and the corrected isenthrops are about parallel, The point C123 is selected close to the saturation curve at the condensation temperature t2. The horizontal distance C123-C0 is then according to the invention preferably selected to 1-2% of the latent enthalpy and no more than 5-10%. The said superheating may be made directly in the evaporators and separately superheaters are not necessary. The endpoint from the vapor expansion C123 is so adjacent to the saturation curve that heat recovery (C123 to C0) in a recuperator not is necessary.

The graph FIG. 9 is typical for most refrigerants. Water and ammonia are diverging where a superheating must be done to avoid moisture in the turbine. However, in US 2014/0026574 A1 the entire refrigerant vapor expansion is made in the moisture region C0-C00 (to the left of the saturation curve). The reason is believed to be, to get condensate as a sealing liquid in an expander (turbine) of the screw type that in turn allows lower shaft speed of the expander.

The multistage evaporations in ECT give different enthalpies for a turbine/expander. Efficiency for axial turbines as well as reaction turbines is basically related to a relation between a vapor/steam velocity c and a peripheral velocity u with a typical dependence as shown in FIG. 10a . The curve has here an accentuated peak. However, when efficiency is expressed versus a relation depending on enthalpy as the inverted value of u/c squared (1/(u/c)²), FIG. 10 b, the corresponding curve is rather flat. This will ease the design of a turbine not only at the rated performance but also for variation in performance and for service at off design conditions.

A New Axial Turbine Design

The new action turbine of the impulse type, FIG. 11, has its design preferably with three sets of nozzles 65, 66 and 67 respectively designed for their actual enthalpy drops. A turbine wheel 52 with centerline 53 can have about conventional blades 51. The nozzles are placed radially considering the velocity ratio u/c. The flow through the turbine blades will then be close to optimum.

For the lowest guiding vane velocity c a separate blade passage may be placed at a smaller radius in the disc. For an ORC (Organic Rankine Cycle) this can be done as peripheral velocities u are considerably lower than for a water steam turbine and structural stresses correspondingly lower. Alternatively, a second turbine wheel could be used.

For low mass flow loaded wheels the three nozzle groups may be arranged at an equal wheel radius.

A New Radial Vapor/Steam Turbine for an ECT Process

An ECT radial/mixed flow turbine, FIG. 12, has a rotatable runner 50 with blades 51 connected by a hub 52 to a shaft with a centerline 53. The runner has preferably a stepped outside diameter with the diameters D2 and D2′ respectively. A radially extended part between said diameters may have a separation wall 54. The shaft with bearings carries the overhung runner and shaft seals built in a conventional way.

Said runner is operable in a casing 56 with volute shaped parties a first 61, a second 62 and a third 63 placed at at least on another part of the circumference. They are connected to different inlets (not shown) from the ECT working fluid (process) circuits. In the casing a holder 64 equipped with guiding vanes sets a first one 65, a second one 66 and a third 67 are clamped by a cover 57 connected to the casing by screws 58. Said cover has an outlet opening with a connection flange 59.

The runner blades and the said cover has a relatively tight clearance as well as the radially extended distance between the runner wall 54 and the vane holder 64 in order to avoid mitigation of the turbine performance. The rotatable separation wall 54 may be replaced by said holder 64 extended radially inwards to meet the diameter D2′ and having a tight axial clearance to the blades 51.

The at least two ECT process circuits have different enthalpy drops that in turn at an expansion give different absolute velocities c. For the fluid design of the turbine with a peripheral velocity u and an absolute velocity c just outside the runner, a velocity ratio u/c should equal about one to obtain a high efficiency. The desired velocity ratio u/c is obtained by varying either u or c or even both of them simultaneously.

In the radial vaneless gaps, given by the diameter differences D4−D2 and D4′−D2′ respectively, the flow follows a free vortex law that has the feature; that radius times the tangential component of the absolute velocity c_(u) is constant. A smaller radius in the gap then gives a higher velocity c_(u). Here the velocity c is about equal to c_(u). Then the desired ratio u/c is obtained by selecting the diameters D2, D2′ and/or D4, D4′ accordingly.

The general configuration may be varied in several ways known per se. The semi open runner in FIG. 13a may be almost fully open with a tight clearance to the casing and cover on both sides or closed with a rotatably runner cover connected to the blades. This runner cover may have a sealing part to the casing cover. In order to reduce thrust, a similar sealing could be arranged on the opposite runner side. The casing as shown is radially split with access on the outlet connection side 59. Access could also be arranged on the opposite side. The radially split casing could also be replaced by a casing axially split through the centerline 53.

The velocity ratios u/c are different for an axial impulse turbine (de Laval action type with 0% reaction) with u/c about 0.4-0.5 and a radial turbine with 50% reaction where u/c is equal to about one. For the radial turbine with 50% reaction c is about 70% of the value of c for an action turbine.

For a radial turbine the different radii may be arranged as a stepped outside wheel diameter with 3 steps (the FIG. 12 shows two diameter steps). Alternatively, the radii nozzle bank may have their outlets on different radii with a different radial gap to the turbine disc. In this gap then we get a free vortex gap that is designed to fit the entrance flow to the wheel.

ECT Applied on Some Other Turbine/Expander Machine Types.

The basic principle of the ECT can be applied on other expander/turbine types as on the broad family of positive displacement machines. As an easily understandable example a triple (3 stages/cylinders in series) compound steam/vapor machine shall be mentioned. The adoption to ECT means that part working fluid streams are added between the stages and the cylinder sizes are adapted to the corresponding increased vapor flows.

The screw expander with 2 screws, developed from the Lysholm screw compressor, or a single screw expander with side mounted sealing wheels can also be designed as a 2 or 3 stage machine with intermediate inlets for adding part flow streams between the stages. The screw volume capacity is changed by selecting the size and/or the number of the “pistons” in the male and the slots in the female rotor/rotors according to the volume flow. Note, when these changes are arranged within common rotors, the screw pitch must be equal for all parts.

ECT Applications

Converting heat to cost efficient electricity generated at:

-   -   Industrial processes     -   Geothermal heat at low temperature wells     -   Solar thermal heat especially with hot water storage for 24         hours service     -   Seasonal heat surplus in district heating.     -   Bottoming cycles for gas turbines, IC engines and power plants.

By ECT the electric output/gain is more than doubled compared to conventional commercial ORC processes using the same heat source and the same heat sink. ECT compares also well with non-isothermal vaporization processes. In addition ECT has a simpler structure and better conditions for the turbine/expander.

The ECT improvement is based on a systematic thermodynamical analysis, first defining a theoretical process analogous with the Carnot process working between two constant temperatures and then applying the result to practical conditions regarding working fluid and process components. 

1. An apparatus of Organic Rankine Cycle type including a closed loop working fluid circuit (2) operable between a heat source (1) and a heat sink (3), the working fluid circuit including: a heat exchanger arrangement (12-14) for vaporizing and/or superheating a working fluid by exchanging energy from the heat source (1); at least one turbine (15) for expanding the vaporized/superheated working fluid; condensing means (16) connectable to the heat sink (3) for condensing the expanded working fluid from the turbine; and pumping means (11) for pumping and pressurizing the condensed working fluid to the heat exchanger arrangement (12-14); wherein the heat exchanger arrangement (12-14) comprising at least three parallel coupled evaporators forming at least three pressure stages (L, M, H) on the working fluid side, and providing one outlet per pressure stage for connecting to the at least one turbine (15).
 2. The apparatus according to claim 1, wherein the at least one turbine (15) is a single turbine (15), said turbine including one inlet per pressure stage connecting to respective pressure stage outlet, and a common outlet (59) for the expanded working fluid.
 3. The apparatus claim 1, wherein the heat exchanger arrangement includes: three evaporators, one for a low pressure stage (13L), one for a medium pressure stage (13M), and one for a high pressure stage (13H); at least one economizer (12L, 12M, 12H) per pressure stage; and optionally at least one superheater (14L, 14M, 14H) per pressure stage.
 4. The apparatus claim 1, wherein the heat exchanger arrangement includes: one economizer (12L) at the low pressure stage, two economizers (12M′, 12M) at the medium pressure stage, and three economizers (12H″, 12H′, 12H) at the high pressure stage.
 5. The apparatus claim 1, wherein the heat exchanger arrangement (12-14) is configured to heat the working fluid at each pressure stage (L, M, H) to an individual temperature starting point (S I, S2, S3) at the corresponding turbine inlet, each starting point (S I, S2, S3) selected to be in the dry/superheated region.
 6. The apparatus according to claim 5, wherein the heat exchanger arrangement (12-14) is configured such that each starting point (S I, S2, S3) is selected to provide the turbine expansion to end in within +/−5° C. of a common temperature and pressure end point (CI 23) in the dry/superheated region.
 7. The apparatus according to claim 6, wherein the common end point (C123) is situated about 5-10% of the latent heat/enthalpy (CO to COO) in the dry/superheated region of the saturation curve from the saturation curve.
 8. The apparatus according to claim 7, wherein the common end point (C123) is situated about 1-2% of the latent heat/enthalpy (CO to COO) in the dry/superheated region of the saturation curve from the saturation curve.
 9. The apparatus claim 1, wherein at least two of the evaporators (13L, 13M, 13H) are coupled in series on the heat source side, such that at least a portion of the heat source flow is directed through the evaporators (13L, 13M, 13H) in series.
 10. The apparatus claim 1, wherein the pressure stages are selected such that the vaporization temperatures of the evaporators are within T2+An*(Tl−T2)+/−15%, where An=(l/(n+l), 2/(n+l), . . . , (n)/(n+l)), where n being the number of evaporators, and where Tl being the temperature of the heat source and T2 the temperature of the heat sink.
 11. The apparatus according to claim 2, wherein the turbine (15) is of a turbo machine type being an impulse/action type or a reaction turbine and having at least three inlets, and guiding vanes and/or nozzles (65, 66, 67) for directing a flow from said inlets on to different radii of one common turbine wheel (52) of the turbine.
 12. The apparatus according to claim 11, wherein the turbine is a reaction turbine of a radial or a mixed flow type with a centripetal (inwards) flow direction with a casing (56) and at least one runner/wheel (50) with working blades (51), at least three inlets and one common outlet (59) where the inlets in turn are connectable to guiding vanes and/or nozzles (65, 66, 67) in the turbine casing for expansion of the gas at different radii.
 13. The apparatus according to claim 11, wherein the turbine is an impulse/action type with a disc (52) that has blades (51) at different radii to match the driving expanded flow from the guiding vanes and/or nozzles (65, 66 and 67) whereby at least two sets of the vanes/nozzles are situated at different radii and the remaining at equal radii but at separate angular sectors.
 14. The apparatus according to claim 12, wherein having at least two of the said guiding vanes and/or nozzles situated on different diameters D4 and D4′ respectively.
 15. The apparatus according to claim 14, wherein having the said runner working blades with an extension to at least two different diameters D2 and D2′ respectively.
 16. The apparatus according to claim 1, wherein the turbine is of a positive displacement machine type with several stages by having displacement volumes determined by a first actual flow stream of the working fluid to which part working fluid streams are further added between the stages and that the downstream displacement sizes are correspondingly increased to swallow the total vapor flows.
 17. The apparatus according to claim 16, wherein the displacement machine includes one or several screws designed to be a 2 or 3 stage machine with intermediate inlets for said part flow streams between the stages and that the downstream screw displacement/volume capacity is designed by selecting shape, size and/or the number of the pistons in a first/male and slots in a second/female rotor cooperating with the first to swallow both the actual and the total volume flows. 